First and referring to FIG. 1, a known refrigeration system 10 for a heat pump, refrigerator, chiller or air conditioner is shown schematically for background purposes. The known refrigeration system 10 includes a compressor 11, driven by an electric motor 12 or other known means, that compresses vapor. The compressor 11 discharges compressed vapor, at high pressure and high temperature, into a condenser 13 where heat is extracted from the working fluid, causing condensation of the high pressure vapor into high pressure liquid. The high pressure liquid then flows from the condenser 13 into a throttling valve 14 which reduces the pressure of the liquid, causing partial flashing. This lower pressure fluid is then routed into an evaporator 15 in which the fluid absorbs heat, thereby converting the working fluid from the liquid to the vapor state. The vapor from the evaporator reenters the compressor 11 on the intake side.
FIG. 2 shows a vapor compression cycle PH (pressure v. enthalpy) diagram for the conventional refrigeration system shown in FIG. 1. with pressure (P) represented along the ordinate and enthalphy (H) appearing along the abscissa. The vapor/compression cycle shows an adiabatic compression of vapor along line A, superheated cooling of the vapor occurring along line B1, followed by biphase isothermal condensation along line B2, and liquid subcooling along line B3. When the working fluid passes through a throttling valve, the working fluid undergoes isoenthalpic expansion, as indicated by vertical line C. Isobaric evaporation of the liquid in the evaporator is shown by horizontal line D.
As should be apparent from the preceding diagram, and with isoenthalpic expansion, the quality of the expanded refrigerant is increased because some of the compression energy of the condensed working fluid is consumed in transforming the liquid into vapor at the low pressure side of the system. For efficient operation, the quality of the working fluid; that is, the vapor fraction of the expanded refrigerant, should be as small as possible.
Referring to FIG. 3, an improved system has been developed, as described in commonly owned U.S. Pat. No. 5,467,613, in which a turbine expander 17 is substituted for the throttling valve expander. The turbine expander 17 receives the high pressure liquid from the condenser and drives a turbine rotor with the kinetic energy of the expanding working fluid. In other words, a portion of the energy imparted to the working fluid by the compressor is recovered in the expander as mechanical energy. Therefore, the turbine expander relieves some of the compressor load on the drive motor, so that the refrigeration cycle operates more efficiently than is possible with a throttling type of expander.
Typically, the turbine expander is either mechanically or electrically connected with the main compressor. A typical mechanical arrangement is illustrated in FIG. 3. A disadvantage of the direct coupling arrangement is that the turbine/expander must be placed in close proximity with the main compressor. This results in the need for additional piping in the system and consequently increases the implementation cost of the two-phase flow expander.
Another possible solution to the above problem, shown in FIG. 4, is to provide a stand alone turbine/expander which locally transfers its recovered mechanical power into electrical power through the use of a generator 18. This transferred electrical power supplies a portion of the electrical power that is required to drive the motor 12 of the compressor 11. The disadvantage with this system is the need for the additional electric generator, as well as the additional losses associated with the generator.
In addition, each of the systems shown in FIGS. 3 and 4 require turbine/expanders which are run at fixed speeds. In actual system applications, however, fixed speed operation requires additional hardware to prevent hot gas by-pass from the condenser to the evaporator during part load conditions. As a consequence, the efficiency of existing throttle loss recovery systems deteriorates under part-load conditions. For example, for a system running at or below 50% capacity with reduced temperature lift, it has been found that power recovery of the turbine/expander is typically reduced to almost negligible amounts.